Control of in-cylinder conditions of an internal combustion engine operating with multiple combustion modes

ABSTRACT

A method of controlling a diesel engine that is capable of multiple combustion modes and equipped with a turbocharger and EGR loop. The control method avoids a singularity condition inherent in turbocharged diesel engine having multiple combustion modes. For different combustion modes, different system states, control variables, and actuators are carefully chosen for different controllers based on the characteristics of the corresponding combustion mode as well as sensor and measurement limitations.

RELATED APPLICATION

This application is a continuation of U.S. patent application Ser. No.11/691,720 filed Mar. 27, 2007, the contents of which is herebyincorporated in its entirety by reference.

TECHNICAL FIELD OF THE INVENTION

This invention relates to control systems for internal combustionengines, and more particularly to a control system for an internalcombustion engine having multiple combustion modes.

BACKGROUND OF THE INVENTION

In recent years, it has become apparent that conventional dieselcombustion cannot alone meet emission levels mandated for the future.Hence, diesel engine manufacturers have been considering multiplecombustion modes as a means to reduce emissions. Alternate combustionmodes such as homogeneous charge compression ignition (HCCI), lowtemperature combustion (LTC), and premixed charge compression ignition(PCCI) are being developed and implemented on diesel engines, togetherwith conventional diesel combustion.

At steady-state, alternate combustion modes offer great potential toreduce engine emission levels. However, because the applicablespeed-load regions of different combustion modes are different from eachother, the engine must seamlessly switch among these modes.

The different combustion modes are achieved by different fueling andin-cylinder conditions. Some modes are close to the edge of unstablecombustion, and are very sensitive to engine conditions.

For diesel engines, fueling control can be exercised precisely on acycle-by-cycle basis. However, in-cylinder conditions change at a muchslower rate (over several combustion cycles). Poor control overin-cylinder conditions not only diminishes the merits of alternatecombustion modes but also worsens drivability and emissions.

BRIEF DESCRIPTION OF THE DRAWINGS

A more complete understanding of the present embodiments and advantagesthereof may be acquired by referring to the following description takenin conjunction with the accompanying drawings, in which like referencenumbers indicate like features, and wherein:

FIG. 1 illustrates a diesel engine suitable for multiple combustionmodes, modeled in accordance with the invention.

FIG. 2 illustrates a finite state machine representing mode switchingbetween a conventional combustion mode and a low temperature combustionmode.

FIG. 3 illustrates the switching surface between the conventionalcombustion mode and the low temperature combustion mode.

FIG. 4 is a block diagram of a control system for a diesel engine havingmultiple combustion modes.

FIGS. 5A and 5B illustrate transient torque responses and exhaust AFRfor transitioning from low temperature combustion to conventionalcombustion in accordance with a conventional control method.

FIGS. 6A and 6B illustrate transient torque responses and exhaust AFRfor transitioning from low temperature combustion to conventionalcombustion in accordance with the invention.

DETAILED DESCRIPTION OF THE INVENTION Dynamic Engine Model as Basis forControl System

One approach to designing control systems for engine in-cylinderconditions is to first develop a dynamic model of the engine. The modelcan be implemented in a graphical simulation environment, using real orsimulated engine conditions as inputs. Appropriate software is used tobuild the model and to manage data.

Once the model is developed, additional software can be used to developan engine controller that can accept various engine conditions andachieve desired performance. After developing and tuning the controlsystem through simulations, control system programming is easilygenerated for production-level controller hardware. In sum, good enginemodels are indispensable for a transition from a rapid prototypingcontroller to a production-level controller.

For purposes of this invention, the modeling is of the intake andexhaust of an engine operating with multiple combustion modes. Thedynamics of various intake and exhaust passage sections and componentsof the engine are modeled, based on physical laws with intentionalsimplifications and reductions. The resulting dynamic model is used asthe basis for design of an in-cylinder engine condition control system.

A feature of the model is that it models only those enginecharacteristics necessary for multiple mode engine control. That is, itdoes not attempt to comprehensively model the entire engine. Varioussimplifications are made, such as modeling the turbocharger dynamics asa first order system. The model thereby simplifies the process ofdesigning a control system suitable for controlling multiple engine modetransitions.

Dynamic Engine Modeling System

FIG. 1 illustrates an engine 100, capable of operating with multiplecombustion modes. An example of such an engine 100 is a light duty4-cylinder common rail diesel engine. The engine is equipped with aturbocharger 110, and a high pressure EGR loop 120 with two paths(cooled and uncooled). The tailpipe has various exhaust treatmentdevices, such as a diesel oxidation catalyst 171, diesel particulatefilter 172, and lean NOx trap 173.

Modeling system 30 models engine 10 for purposes of designing a controlsystem 20 for operation of, and transitioning between, multiple enginemodes. As explained below, modeling system 30 is used to determine howvarious actuators can be controlled to provide desired pressure and airfraction conditions of engine 10. Modeling system 30 can be implementedwith computer equipment programmed to store and execute the equationsand data described herein. As explained above, control system 20 isdesigned using modeling system 30, and for production engines isimplemented with appropriate engine control system hardware andsoftware.

A more complete description of modeling system 30 is set out in U.S.Patent Application Ser. No. 60/836,818 (Atty Dkt No. 090936.0699),entitled “Dynamic Modeling of an Internal Combustion Engine Operatingwith Multiple Combustion Modes”, incorporated by reference herein.

Various sections associated with the air intake and exhaust of engine100 are each represented in the model. These include:

Section 1 intake manifold (between the high-pressure throttle and high-pressure EGR valve and engine intake valves) Section 2 intake passagebetween compressor and high pressure throttle Section 3 exhaust manifoldSection 4 intake to turbocharger

As compared to a model for an engine having only conventionalcombustion, a model of engine 100 requires more variables and systemstates. In general, evolution of in-cylinder conditions can be viewed asa multi-variable nonlinear system.

The model described herein may be generally described as an “engineintake and exhaust system dynamic model”. The actuators of interest forthe model are the turbocharger 110 (its output flow), the intakemanifold throttle 130, and EGR throttle 150. The model is especiallydirected to the dynamics of the fresh air fraction in the intakemanifold, that is, the ratio of fresh air from the compressor to theamount of recirculated exhaust gas, and the various pressures.

The actuators are used to control the fresh air fraction and pressure sothat these parameters are appropriate for a given combustion mode. If adesired pressure or fresh air fraction is known, model system 30 can beused to determine how to actuate throttles 130 and 150 and the output ofturbocharger 110.

Turbocharger 110 has a compressor 111 and turbine 112, and is assumed tobe a variable output turbocharger. An example of a suitable turbochargeris a variable geometry turbocharger (VGT). As explained below, thecompressor power is modeled as a differential equation.

As stated above, engine 100 also has an EGR (exhaust gas recirculationloop), which is a high pressure loop. EGR cooler 121 cools the exhaustbefore it is mixed with fresh air from the compressor 111.

Temperatures at various intake and exhaust points of engine 100 are alsorepresented in the model. FIG. 1 illustrates the location of varioustemperature and pressure measurement sensors, for sensing T1, T2, and T3(temperatures) and P1, P2, and P3 (pressures). Some temperature andpressure values can be inferred or assumed. For example, P4 is assumedto be the atmospheric pressure. T5 can be inferred from T2 and theintercooler efficiency. An O2 sensor 174 is installed to measure the O2in the exhaust from the exhaust manifold.

Engine 100 has several intake or exhaust sections, which are labeled 1-4in FIG. 1. The intake and exhaust modeling is based on principles ofmass and energy conservation as well as the ideal gas law.

Pressures in the various sections can be expressed as:

$\begin{matrix}{{{\overset{.}{p}}_{1} = {{{- \frac{\gamma \; \eta_{v}N_{e}V_{d}}{120\; V_{1}}}p_{1}} + {\frac{R\; \gamma \; T_{HT}}{V_{1}}W_{HT}} + {\frac{R\; \gamma \; T_{EGR}}{V_{1}}W_{EGR}}}}{{\overset{.}{p}}_{2} = {\frac{R\; \gamma \; T_{c}\eta_{c}P_{c}}{V_{2}C_{p}{T_{4}\left\lbrack {\left( \frac{p_{2}}{p_{4}} \right)^{\frac{\gamma - 1}{\gamma \;}} - 1} \right\rbrack}} - {\frac{R\; \gamma \; T_{HT}}{V_{2}}W_{HT}}}}{{\overset{.}{p}}_{3} = {\frac{R\; \gamma}{V_{3}}\left\lbrack {{\frac{T_{eo}\eta_{v}N_{e}V_{d}}{120\; {RT}_{1}}p_{1}} + {T_{eo}W_{f}} - {T_{3}W_{EGR}} - {T_{3}W_{t}}} \right\rbrack}}{{\overset{.}{P}}_{c} = {{- \frac{P_{c}}{\tau_{tc}}} + {{\frac{C_{p}T_{3}\eta_{t}}{\tau_{tc}}\left\lbrack {1 - \left( \frac{p_{t}}{p_{3}} \right)^{\frac{\gamma - 1}{\gamma}}} \right\rbrack}W_{t}}}}} & (1)\end{matrix}$

For a diesel engine, combustion is usually lean, which means that theair in the cylinder mixture exceeds the stoichiometric amount. As aresult, the exhaust gas contains unburned air and can be re-circulatedback into the intake manifold through EGR valve 150. The fraction of air(or EGR gas) in-cylinder is important for combustion and emissionsperformance, especially for alternate combustion modes, which can beclose to unstable.

The dynamics of the fresh air conditions can be described as:

$\begin{matrix}{{\overset{.}{F}}_{1} = {\frac{{RT}_{1}}{p_{1}V_{1}}\left\lbrack {{\left( {1 - F_{1}} \right)W_{HT}} + {\left( {F_{3} - F_{1}} \right)W_{EGR}}} \right\rbrack}} & \left( {2a} \right) \\{{\overset{.}{F}}_{3} = {\frac{{RT}_{3}}{p_{3}V_{3}}{W_{eo}\left( {F_{eo} - F_{3}} \right)}}} & \left( {2b} \right) \\{F_{eo} = {\frac{{W_{e}F_{1}} - {W_{f}\lambda_{s}}}{W_{e} + W_{f}} = \frac{W_{f}\left( {\lambda_{m} - \lambda_{s}} \right)}{W_{e} + W_{f}}}} & \left( {2c} \right) \\{F_{1} = \frac{W_{f}\lambda_{m}}{W_{e}}} & \left( {2d} \right)\end{matrix}$

In the above equations, λ_(s) is the stoichiometric air to fuel ratio,λ_(m) is the measured air to fuel ratio by a UEGO (universal exhaust gasoxygen) sensor installed on the exhaust manifold. The value F₃ can beestimated from Equation (2b) as:

$\begin{matrix}{F_{3} = {\frac{\frac{{RT}_{3}}{p_{3}V_{3}}W_{eo}}{s + {\frac{{RT}_{3}}{p_{3}V_{3}}W_{eo}}}F_{eo}}} & (3)\end{matrix}$

If the pressure-drop of the inlet air filter is ignored, then thecondition in Section 4 of the engine is same as atmosphere, p₄=p_(a),T₄=T_(a). The value p_(a) is atmosphere pressure, which is assumed knownand can be measured by standard sensors. The value T_(a) is the inletair temperature, which is available from mass airflow (MAF) sensor 140.

For modeling system 30, the system states are x₁=p₁, x₂=p₂, x₃=p₃,x₄=P_(c), and x₅=F₁, or x=[p₁ p₂ p₃ P_(c) F₁]^(T). The values p₁, p₂ andp₃ can be measured by sensors as shown in FIG. 1. The values P_(c) andF₁ can be calculated or estimated based on other measurements such asMAF, AFR, and temperatures.

The system inputs are u₁=W_(HT), u₂=W_(EGR), and u₃=W_(t), or u=[W_(HT)W_(EGR) W_(t)]^(T), corresponding to gas flow rates through thehigh-pressure throttle, EGR valve, and turbine, respectively. These flowrates can be converted into angles of the high pressure throttle valve130, the EGR valve 150, and the nozzle position of the variable geometryturbocharger (VGT) 110, through inverse orifice equations or mappings.

The fueling rate W_(f) is determined by engine maps and treated as anexternal signal. The value p_(t) is the turbine pressure, which can becalculated based on the signal of the exhaust section delta pressuresensor as p_(t)=p_(a)+Δp_(ex).

For notation simplicity, let

${k_{1} = \frac{\gamma \; \eta_{v}N_{e}V_{d}}{120\; V_{1}}},{k_{2} = \frac{R\; \gamma \; T_{c}\eta_{c}p_{a}^{{({\gamma - 1})}/\gamma}}{V_{2}C_{p}T_{a}}},{k_{{ht}\; 1} = \frac{R\; \gamma \; T_{HT}}{V_{1}}},{k_{r\; 1} = \frac{R\; \gamma \; T_{EGR}}{V_{1}}},{k_{{ht}\; 2} = \frac{R\; \gamma \; T_{HT}}{V_{2}}},{k_{3} = \frac{\gamma \; T_{eo}\eta_{v}N_{e}V_{d}}{120\; V_{3}T_{1}}},{k_{r\; 3} = {k_{t\; 3} = \frac{R\; \gamma \; T_{3}}{V_{3}}}},{k_{f} = \frac{R\; \gamma \; T_{eo}}{V_{3}}},{k_{t\; 1} = \frac{C_{p}T_{3}\eta_{t}}{\tau_{tc}}},{k_{t\; 2} = \frac{C_{p}T_{3}{\eta_{t}\left( {p_{a} + {\Delta \; p_{ex}}} \right)}^{{({\gamma - 1})}/\gamma}}{\tau_{tc}}},{k_{4} = \frac{{RT}_{1}}{p_{1}V_{1}}},{and}$$\kappa = {\frac{\gamma - 1}{\gamma}.}$

Thus, the engine model used by modeling system 30 can be rewritten as:

$\begin{matrix}{{{\overset{.}{x}}_{1} = {{{- k_{1}}x_{1}} + {k_{{ht}\; 1}u_{1}} + {k_{r\; 1}u_{2}}}}{{\overset{.}{x}}_{2} = {{k_{2}\left( \frac{x_{4}}{x_{2}^{\kappa} - p_{a}^{\kappa}} \right)} - {k_{{ht}\; 2}u_{1}}}}{{\overset{.}{x}}_{3} = {{k_{3}x_{1}} - {k_{r\; 3}u_{2}} - {k_{t\; 3}u_{3}} + {k_{f}W_{f}}}}{{\overset{.}{x}}_{4} = {{{- \frac{1}{\tau_{tc}}}x_{4}} + {k_{t\; 1}u_{3}} - \frac{k_{t\; 2}u_{3}}{x_{3}^{\kappa}}}}{{\overset{.}{x}}_{5} = {{{k_{4}\left( {1 - x_{5}} \right)}u_{1}} + {{k_{4}\left( {F_{3} - x_{5}} \right)}u_{2}}}}} & (4)\end{matrix}$

In the above equations, values are obtained for pressure at threesections of the engine as well as for the compressor power and a freshair value. These values are used to model the behavior of engine 100.Using the model, control unit 20 can be programmed to controlin-cylinder conditions for optimum engine control, including determiningwhen to switch modes and conditions that will achieve optimum modetransitions.

More specifically, for a given combustion mode, certain parameters suchas pressure and fresh air ratio are desired. The model can be used todetermine which engine inputs will result in desired outputs. The inputsinclude the positions of the HP throttle 130 and EGR throttle 150 andthe flow through turbine 112.

The above-described modeling approach (modeling the dynamics of eachsection in the intake and exhaust passages with necessary measurementand estimations) can be easily expanded for engines with differentintake and exhaust system configurations, such as dual-loop EGR systems.

Singularity Issues

The above-described model can be expressed in a general state spaceform, which is:

$\underset{\underset{x}{}}{\begin{pmatrix}{\overset{.}{x}}_{1} \\{\overset{.}{x}}_{2} \\{\overset{.}{x}}_{3} \\{\overset{.}{x}}_{4} \\{\overset{.}{x}}_{5}\end{pmatrix}} = {\underset{\underset{f{(x)}}{}}{\begin{pmatrix}{{- k_{1}}x_{1}} \\\frac{k_{2}x_{4}}{x_{2}^{\kappa} - p_{a}^{\kappa}} \\{{k_{3}x_{1}} + {k_{f}W_{f}}} \\{- \frac{x_{4}}{\tau_{tc}}} \\0\end{pmatrix}} + {\underset{\underset{g{(x)}}{}}{\begin{pmatrix}k_{{ht}\; 1} & k_{r\; 1} & 0 \\{- k_{{ht}\; 2}} & 0 & 0 \\0 & {- k_{r\; 3}} & {- k_{t\; 3}} \\0 & 0 & {k_{t\; 1} - \frac{k_{t\; 2}}{x_{3}^{\kappa}}} \\{k_{4}\left( {1 - x_{5}} \right)} & {k_{4}\left( {F_{3} - x_{5}} \right)} & 0\end{pmatrix}}\underset{\underset{u}{}}{\begin{pmatrix}u_{1} \\u_{2} \\u_{3}\end{pmatrix}}}}$ $\overset{.}{x} = {{f(x)} = {{g(x)}u}}$

The engine, as modeled above, has a singularity at x₂=p₂=p_(a) whichmeans compressor flow becomes infinite. For conventional controlsystems, it is claimed that p₂ will be above atmosphere pressure all thetime. However, for a turbocharged diesel engine running alternatecombustion modes, such as LTC, the intake manifold pressure could varyfrom much lower than atmosphere pressure to much higher than atmospherepressure. Therefore, this singularity problem needs to be addressed.

In addition, several parameters in the model are uncertain and there arealso some un-modeled dynamics in the actual system. Therefore, thecontrol system has to be sufficiently robust to handle the parameteruncertainties and un-modeled dynamics.

Control System

As stated above, different combustion modes have differentcharacteristics and operating ranges. System outputs for control system20 to track the desired values are selected for different combustionmodes. In this example of this description, two combustion modes, lowtemperature combustion (LTC) and conventional diesel combustion, areconsidered. In the following equations, to distinguish the differentcontrollers and variables, the subscripts “a” and “c” are used for thealternate (LTC) and conventional combustion modes respectively.

Control for LTC Combustion Mode

Combustion features of the LTC mode are: low fresh airflow rate, highEGR rate, low AFR (about 17˜21), low intake manifold pressure (about70˜98 kPa), and low torque range (about 0˜60 Nm).

To perform good closed-loop control, measurements or estimations ofcertain system signals need to be reliable and accurate. In particular,intake fresh air, EGR amounts, and pressure measurements or estimationsare important.

However, in the LTC mode, the noise-to-signal ratio of the MAF sensorsignal is high, and measurement accuracy can be poor due to low freshairflow rate. The MAF measurement may not be reliable enough forclosed-loop control purposes. On the other hand, because the exhaust AFRis low (close to stoichiometric), the measured AFR ratio signal from atypical UEGO sensor is accurate and reliable. Thus, the estimated intakemanifold fresh air fraction, F₁, calculated from Equation (2d) usingmeasured AFR is reliable.

Therefore, y_(a)=[p₁ F₁]^(T) is the system output vector that controller20 will use to manipulate actuators to track the desired valuesy_(ad)=[p_(1d) F_(1d)]^(T).

Because the intake manifold pressure is lower than atmosphere pressurefor LTC, the turbocharger is not used in this mode. The actuators arethen the high-pressure throttle valve 130 and EGR valve 150. The systemstate-space equations are:

$\begin{matrix}{{{\overset{.}{y}}_{a} = {\begin{pmatrix}{{- k_{1}}x_{1}} \\0\end{pmatrix} + {\begin{pmatrix}k_{{ht}\; 1} & k_{r\; 1} \\{k_{4}\left( {1 - x_{5}} \right)} & {k_{4}\left( {F_{3} - x_{5}} \right)}\end{pmatrix}\begin{pmatrix}u_{1} \\u_{2}\end{pmatrix}}}}{u_{a} = {f_{a}\left( {y_{a},y_{ad},x,k_{a}} \right)}}} & (7)\end{matrix}$

By tracking the desired [p_(1d) F_(1d)]^(T), the fresh air amountin-cylinder will be close to the desired value. Several advanced controlsystem design methods such as sliding mode control, feedbacklinearization, and Lyapunov-based control design can be used for thetracking control design. For experimental test results, amulti-input-multi-output sliding mode control was designed for thissystem. It could be shown that the zero dynamics involved for the restof the system state is stable.

Control for Conventional Diesel Combustion Mode

Some of the combustion features of the conventional combustion mode are:higher fresh airflow rate, relatively lower EGR rate, higher AFR (about21˜28), higher intake manifold pressure (above atmosphere pressure), andhigher torque range (60 Nm up to peak torque).

To control combustion at conventional diesel combustion mode, intakemanifold pressure, fresh air charge/EGR rate are important variables.Control system 20 needs to track the desired values.

In the conventional combustion mode, the exhaust AFR is higher than inLTC mode, and the signal from a typical UEGO sensor may not be accurate.On the other hand, since the fresh air mass flow rate is higher than inLTC mode, a typical production MAF sensor should provide a sufficientlyaccurate measurement.

Based on these considerations, the system output vector is y_(c)=[p₁W_(c) p₃]^(T). The value W_(c) is the fresh air flow rate through thecompressor. If the control system 20 can track these three variables,the engine intake gas charge amount and EGR rate can be controlled asdesired. The actuators (the available control inputs) are thehigh-pressure throttle, EGR valve, and VGT.

$\begin{matrix}{W_{C} = {\frac{P_{c}\eta_{c}}{C_{p}{T_{4}\left\lbrack {\left( \frac{p_{2}}{p_{4}} \right)^{\frac{\gamma - 1}{\gamma}} - 1} \right\rbrack}} = \frac{P_{c}\eta_{c}}{C_{p}{T_{a}\left\lbrack {\left( \frac{p_{2}}{p_{a}} \right)^{\kappa} - 1} \right\rbrack}}}} & (8) \\{\begin{matrix}{{\overset{.}{W}}_{C} = {\frac{\eta_{c}{\overset{.}{P}}_{c}}{C_{p}{T_{a}\left\lbrack {\left( \frac{p_{2}}{p_{a}} \right)^{\kappa} - 1} \right\rbrack}} + \frac{{- P_{c}}{\eta_{c}\left( {\kappa \; \frac{p_{2}^{\kappa - 1}}{p_{a}^{\kappa}}} \right)}{\overset{.}{p}}_{2}}{C_{p}{T_{a}\left\lbrack {\left( \frac{p_{2}}{p_{a}} \right)^{\kappa} - 1} \right\rbrack}^{2}}}} \\{= {{- \frac{W_{c}}{\tau_{tc}}} + {{c(x)}u_{3}} - {{d(x)}k_{2\; c}W_{c}^{2}} + {{d(x)}W_{c}k_{{ht}\; 2}u_{1}}}}\end{matrix}{{{{Where}\mspace{14mu} {c(x)}} = \frac{\eta_{c}T_{3}{\eta_{t}\left\lbrack {1 - \left( \frac{p_{t}}{x_{3}} \right)^{\kappa}} \right\rbrack}}{\tau_{tc}{T_{a}\left\lbrack {\left( \frac{x_{2}}{p_{a}} \right)^{\kappa} - 1} \right\rbrack}}},{{d(x)} = \frac{\kappa \; p_{2}^{\kappa - 1}}{p_{2}^{\kappa} - p_{a}^{\kappa}}},{{{and}\mspace{14mu} k_{2\; c}} = \frac{R\; \gamma \; T_{c}}{V_{2}}}}} & (9) \\{{{\overset{.}{y}}_{c} = {\underset{\underset{a{(x)}}{}}{\begin{pmatrix}{{- k_{1}}x_{1}} \\{{- \frac{W_{c}}{\tau_{tc}}} - {{d(x)}k_{2\; c}W_{c}^{2}}} \\{{k_{3}x_{1}} + {k_{f}W_{f}}}\end{pmatrix}} + {\underset{\underset{b{(x)}}{}}{\begin{pmatrix}k_{{ht}\; 1} & k_{r\; 1} & 0 \\{{d(x)}W_{c}k_{{ht}\; 2}} & 0 & {c(x)} \\0 & {- k_{r\; 3}} & {- k_{t\; 3}}\end{pmatrix}}\underset{\underset{u_{c}}{}}{\begin{pmatrix}u_{1} \\u_{2} \\u_{3}\end{pmatrix}}}}}{u_{c} = {f_{c}\left( {y_{c},y_{c\; d},x,k_{c}} \right)}}} & (10)\end{matrix}$

By tracking the desired y_(cd)=[p_(1d) W_(cd) p_(3d)]^(T), combustion iswell controlled under transient conditions.

Several advanced control system design methods such as sliding modecontrol, feedback linearization, and Lyapunov-based control design canbe used for the tracking control design. For experimental test results,a multi-input-multi-output sliding mode control system was designed forthis system. It is also easy to show that the zero dynamics involvingthe rest system states is stable.

Supervisory Controller

As explained below in connection with FIG. 4, control system has asupervisory controller whose task is to switch among differentcontrollers based on driver demand and engine operating conditions. Fora turbocharged diesel engine, another purpose of the supervisorycontroller is to avoid the above-described singularity condition.

The singularity occurs when p₂ is close to p_(a) and turbocharger 110 isused, and must be avoided to prevent undesired behaviors of controlsystem 20. In LTC mode, the intake manifold pressure is below p_(a) andturbocharger 110 is not used. However, when the intake manifold pressureapproaches p_(a), the system is close to the switching surface from LTCto conventional diesel combustion.

To prevent the singularity condition happening at conventional dieselcombustion, the VGT is intentionally offset to increase p₂ above p_(a).Thus, p₂ is above p_(a) from the beginning of the conventional dieselcombustion mode. When LTC engine operation conditions are close to theswitching surface, the intake manifold pressure is close to p_(a) andthe exhaust gas has sufficient energy to push the turbine to increase p₂by the compressor. As p₂ is increased at the high end of LTC combustion,the high-pressure throttle is automatically adjusted to control p₁ asdesired.

FIG. 2 illustrates a finite state machine (FSM) for conducting the modeswitching task. As illustrated, the FSM has three states:

1. Conventional Combustion Controller

2. LTC controller with VGT offset

3. LTC controller without VGT offset.

FIG. 3 illustrates the switching surface for LTC and conventional dieselcombustion modes. The value p_(thd) is the thickness of the switchingsurface from below, and p_(thu) is the thickness of the switchingsurface from above. These values are each a function of engine operatingconditions, such as engine speed.

The value p_(1c) is defined as a function of the desired intake manifoldpressure, p_(1d), and the actual intake manifold pressure, p₁, suchthat:

p _(1c) =f(p _(1d) ,p ₁)  (11)

Coming from the LTC area (Mode 3 in FIG. 2) where turbocharger is notused, if p_(1c)≧p_(s)−p_(thd)+p_(hys), the system is switched to Mode 2where VGT is offset to increase p₂ above p_(a) to prepare switching upto conventional diesel combustion. The value p_(s) is the switchingpressure, which is approximately equal to atmosphere pressure, andp_(hys) is a hysteresis gap to avoid frequent mode switching. At Mode 3,the LTC controller described above is used.

At Mode 2, if p_(1c)≦p_(s)−p_(thd)−p_(hys), it is switched back to Mode3 and VGT goes back to rest position. If p_(1c)≧p_(s)+p_(thu)+p_(hys),it is switched to Mode 1 for conventional diesel combustion. At Mode 2,the LTC controller described above is used.

At Mode 1, if p_(1c)≦p_(s)+p_(thu)−p_(hys), it is switched back to Mode2 for LTC with VGT offset. At Mode 1, the conventional diesel combustioncontroller described above is used.

Overall System Control

FIG. 4 is a block diagram of the overall control system 20. Threevectors of variables in FIG. 4 are: Y, a vector of the system state,Y=[P1, F1, Wc, . . . ]; G, a vector of gas (i.e., air and EGR)actuators, G=[Throttle, EGR, VGT, SCV, VVA, . . . ]; and F, a vector offueling parameters, F=[injection timing, injection quantity, injectionpressure, . . . ]. As explained above, the variables used for thesevectors depend on the combustion mode.

Processing unit 41 receives values for engine speed and the driver'spedal position input, and calculates a desired engine torque.

Processing unit 42 receives engine speed and desired torque, anddetermines a desired engine operation state vector, Y*. This vectorconsists of variables such as intake manifold pressure, fresh airfraction, fresh airflow rate, etc.

The desired state vector, Y*, the measured state vector, Y, and enginespeed are fed to the supervisory controller 43, which determines whichmode and corresponding mode controller should be active.

Processing unit 44 receives the decided mode, engine speed and desiredtorque, and determines engine fueling parameters such as injectionpattern, injection quantities, injection timings and injection pressure,etc.

The desired state vector and measured/estimated state vector are fedinto selected nonlinear robust controller 45. The output of thecontroller 45 together with a feed-forward control, which is a functionof engine speed and desired torque, are combined and delivered toactuator controller 46, which controls various actuators such asthrottle, EGR valve, VGT, etc., to make the actual states track thedesired states.

Processing unit 49 receives signals measured from sensors equipped onthe engine and fueling input, and determines the actual engine state.This state data is delivered to processing unit 43 and 45 as describedabove.

Illustrative Engine Test Results

To illustrate the benefits of the control approach described herein,experimental results from a modern light-duty diesel engine wereobtained. Two measures of good engine performance are engine torqueresponse (for driveability) and exhaust gas AFR (for emission amounts).

FIGS. 5A and 5B illustrate transient torque responses and measuredexhaust AFR, using conventional calibration/mapping based control, forthe transition from a light load LTC mode to a conventional combustionmode. During the combustion mode transition, some torque fluctuationsare exhibited. The AFR significantly deviates from a desired valueespecially at combustion mode transitions, which will cause highengine-out emissions. The main reason for these undesired behaviors isthat the actuators' tables and maps are calibrated at steady-state andcannot account for system dynamics during transient conditions.Un-coordinated movements of actuators cause undesired engine in-cylinderconditions, and cause undesired combustions and emissions.

For comparison, FIGS. 6A and 6B illustrate transient engine torque andAFR responses, using the control methods described herein, during lightload LTC and conventional diesel combustion mode transitions. Comparedwith the conventional control approach, obvious improvements areobserved. Both torque and AFR have smooth responses even duringcombustion mode switching. As a result, the engine satisfies bothdriveability and emissions requirements.

The main reason for this good performance is the closed-loop control onimportant engine operation variables chosen with respect to differentcombustion modes. The actuators are automatically manipulated to accountfor system dynamics. Selected operation variables are tracked close todesired values to ensure good responses for torque and emissions AFR.

1-17. (canceled)
 18. A method of selecting a current combustion mode foroperating a diesel engine capable of multiple combustion modes, such asa normal diesel combustion mode and at least one alternate combustionmode, the engine having a turbocharger and an exhaust gas recirculation(EGR) loop, comprising: defining a set of system state variables foreach combustion mode; wherein the system state variables are based onthe characteristics of the associated combustion mode; during operationof the engine, performing the following steps: acquiring a set ofmeasured or estimated current system state values; receiving an rpmvalue, determining a desired torque value based at least in part on therpm value; determining a set of desired system state values based on thedesired torque value and the rpm value; selecting a combustion mode,based on the rpm value, the current system state values, and the desiredsystem state values.
 19. The method of claim 18, wherein the normaldiesel combustion mode has system state variables of: intake manifoldpressure, the fresh air flow rate through the compressor of theturbocharger, and exhaust manifold pressure.
 20. The method of claim 19,wherein the fresh air flow rate is based on a MAF sensor that sensesfresh air through the compressor of the turbocharger.
 21. The method ofclaim 18, wherein one of the alternate combustion modes is a lowtemperature combustion mode having system state variables of: intakemanifold pressure and the intake manifold fresh air fraction.
 22. Themethod of claim 21, wherein the fresh air fraction is based onmeasurement of the air-fuel ratio.
 23. The method of claim 18, whereinthe measured or estimated system state values are at least one or moreof the following measured or estimated values: the intake manifoldpressure, the pressure between a high pressure throttle between theintake manifold and the compressor of the turbocharger, the exhaustmanifold pressure, the compressor power, and a fresh air value.
 24. Themethod of claim 23, wherein the fresh air value is the fresh airfraction in the intake manifold or the fresh air flow rate through thecompressor.
 25. The method of claim 18, wherein during operation of theengine, the method further comprises measuring a pedal position value,and wherein the desired torque is further based at least in part on thepedal position value.
 26. A method of achieving desired in-cylinderconditions of a diesel engine, capable of multiple combustion modes,such as conventional diesel combustion and at least one alternatecombustion mode, the engine having a turbocharger and an exhaust gasrecirculation (EGR) loop, comprising: defining a set of system statevariables for each combustion mode; defining a set of actuators for eachcombustion mode; wherein the system state variables and actuators arebased on the characteristics of the associated combustion mode; duringoperation of the engine, performing the following steps: acquiring a setof measured or estimated current system state values; receiving an rpmvalue, determining a desired torque value based at least in part on therpm value; determining a set of desired system state values based on thedesired torque value and the rpm value; selecting a combustion mode,based on the rpm value, the current system state values, and the desiredsystem state values; using a controller to determine actuator values,based on the selected combustion mode and the desired system statevalues.
 27. The method of claim 26, wherein the diesel combustion modehas system state variables of: intake manifold pressure, the fresh airflow rate through the compressor of the turbocharger, and the exhaustmanifold pressure; and has actuators of: a high pressure throttlebetween the compressor and the intake manifold, an EGR throttle, and avariable output turbocharger.
 28. The method of claim 27, wherein thefresh air flow rate is based on a MAF sensor that senses fresh airthrough the compressor of the turbocharger.
 29. The method of claim 26,wherein one of the alternate combustion modes is a low temperaturecombustion mode having system state variables of: intake manifoldpressure and the intake manifold fresh air fraction; and havingactuators of: a high pressure throttle between the compressor and theintake manifold and an EGR throttle.
 30. The method of claim 29, whereinthe fresh air fraction is based on measurement of the air-fuel ratio.31. The method of claim 26, wherein the measured or estimated systemstate values are at least one or more of the following measured orestimated values: the intake manifold pressure, the pressure between ahigh pressure throttle between the intake manifold and the compressor ofthe turbocharger, the exhaust manifold pressure, the compressor power,and a fresh air value.
 32. The method of claim 31, wherein the fresh airvalue is the fresh air fraction in the intake manifold or the fresh airflow rate through the compressor.
 33. The method of claim 26, whereinthe actuators are at least one or more of the following actuators: anintake throttle between the intake manifold and the compressor of theturbocharger, an EGR valve, and the turbocharger output.
 34. The methodof claim 26, wherein during operation of the engine, the method furthercomprises measuring a pedal position value, and wherein the desiredtorque is further based at least in part on the pedal position value.35. The method of claim 26, wherein the method is further used toprovide combustion control of the engine, and further providing the stepof using a fueling function to receive the rpm value, the desired torquevalue, and the selected combustion mode and to determine fuelingactuator values.
 36. A control system for achieving desired in-cylinderconditions of a diesel engine, capable of multiple combustion modes,such as conventional diesel combustion and at least one alternatecombustion mode, the engine having a turbocharger and an exhaust gasrecirculation (EGR) loop, comprising: a desired torque map for receivingat least an rpm value and for providing a desired torque value based atleast in part thereon; a desired system state map for receiving thedesired torque value and for providing a set of desired system statevalues based at least in part thereon; a supervisory controller forreceiving the rpm value, the desired system state values, and a set ofmeasured or estimated current system state values, and for providing aselected combustion mode based at least in part thereon; wherein themeasured or estimated system state values are at least one or more ofthe following measured or estimated values: the intake manifoldpressure, the pressure between a high pressure throttle between theintake manifold and the compressor of the turbocharger, the exhaustmanifold pressure, the compressor power, and a fresh air value; and afirst controller for receiving the selected combustion mode and thedesired system state values, and for determining a first set of actuatorvalues based at least in part thereon.
 37. The system of claim 36,wherein the diesel combustion mode has system state variables of: intakemanifold pressure, the fresh air flow rate through the compressor of theturbocharger, and the exhaust manifold pressure; and has actuators of: ahigh pressure throttle between the compressor and the intake manifold,an EGR throttle, and a variable output turbocharger.
 38. The system ofclaim 37, wherein the fresh air flow rate is based on a MAF sensor thatsenses fresh air through the compressor of the turbocharger.
 39. Thesystem of claim 36, wherein one of the alternate combustion modes is alow temperature combustion mode having system state variables of: intakemanifold pressure and the intake manifold fresh air fraction; and havingactuators of: a high pressure throttle between the compressor and theintake manifold and an EGR throttle.
 40. The system of claim 39, whereinthe fresh air fraction is based on measurement of the air-fuel ratio.41. The system of claim 36, wherein the fresh air value is the fresh airfraction in the intake manifold or the fresh air flow rate through thecompressor.
 42. The system of claim 36, wherein the actuators are atleast one or more of the following actuators: an intake throttle betweenthe intake manifold and the compressor of the turbocharger, an EGRvalve, and the turbocharger output.
 43. The system of claim 36, whereinthe desired torque is further based at least in part on the pedalposition value.
 44. The system of claim 36, wherein the control systemfurther provides combustion control of the engine, and furthercomprising a fueling function controller that receives the rpm value,the desired torque value, and the selected combustion mode, and providesfueling actuator values based thereon.